Heat exchanger

ABSTRACT

Air passages  18  are formed between surfaces of a plurality of heat transfer plates  12  laid together. A plurality of rib sections  14  extending orthogonal to the air flowing direction A are formed on a surface of the heat transfer plate  12  and projected into the air passage  18.  By shifting the positions of the rib sections  14  to each other in the air flow direction, coolant passages  15, 16  are formed inside the plurality of rib sections  14.  Between the plurality of rib sections  14,  fin sections  17  are formed integral with the heat transfer plate  12  and projected from the plate surface. The fin section  17  has a protruded shape formed by pressing and having a cut portion partially cut a plate thickness of the heat transfer plate  2.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to a heat exchanger, wherein fins areformed integral with a heat transfer plate constituting internalpassages through which flows a heat exchanging fluid and, for example,useful for a vehicle air conditioner.

2. Description of the Related Art

In the prior art and, for example, in Japanese Unexamined PatentPublication No. 11-287580 (first patent document), a heat exchanger hasbeen proposed wherein a plurality of rib sections, constituting internalpassages through which passes a heat exchanging fluid, are formedintegral with a heat transfer plate and operate as turbulence generatorsfor disturbing a straight flow of air stream flowing on the outside ofthe heat transfer plates.

According to this structure, as an air-side heat transfer rate isimproved by forming turbulence in the air stream, it is possible toeliminate fin members such as the corrugated fins in the conventionalfin and tube type heat exchanger. Thus, the heat transfer plate could beproduced solely by press-forming and brazing heat transfer plates.

Also, in Japanese Unexamined Patent Publication No. 2002-147983 (secondpatent document), as shown in FIG. 27, a heat exchanger is proposedwherein a plurality of rib sections 14, for constituting internalpassages 15 for a heat exchanging fluid, are formed integral with a heattransfer plate 12, and a base plate section 13 having a flat surface isformed between the rib sections 14 adjacent to each other and finsections 17, projected toward an air passage 18, are provided on thebase plate section 13. Also, in this second patent document, the finmembers, such as corrugated fins, are not fixed to the heat transferplate 12.

In this prior art, as the plurality of rib sections 14 constituting theinternal passages are disposed at the same positions as seen in the airflowing direction A, the rib sections 14, 14 in the adjacent heattransfer plates 12, 12 are directly opposed to each other whileinterposing the air passage 18.

As a result, an area S1 of the air passage 18 at a position forming thebase plate section 13, that is, at a position forming the fins 17reduces to S2 at a position forming the rib section 14. Accordingly, inthis prior art, the air passage 18 repeats the reduction and enlargementof the cross-sectional area in accordance with whether or not the ribsection 14 exists.

In this regard, in Japanese Unexamined Patent Publication No. 11-287580,while a local heat transfer rate in the vicinity of the rib section isbetter than in the conventional fin and tube type heat exchanger, theair-side heat transfer area becomes insufficient, whereby there may be acase that a necessary heat transfer performance cannot be ensured.

Also, in the heat transfer plate, as the base plate section having norib section forms a flat surface extending in the air-flow direction, atemperature boundary layer grows on this flat surface to lower the localheat transfer rate to a great extent. This is also a cause for thedeterioration of heat transfer performance.

To ensure the necessary heat transfer performance, it is necessary toincrease the number of heat transfer plates. However, as the heattransfer plate has a large wall thickness for maintaining the requiredpressure resistance in comparison with the fin member, the total weightof the heat exchanger increases. Also, as the material cost of the heattransfer plate increases, the production cost of the heat exchangerbecomes high.

In the second patent document, as the air passage 18 repeats thereduction and enlargement of the cross-sectional area in accordance withwhether or not the rib section 14 exists, a pressure loss in theair-stream increases.

According to the second patent document, as gaps between tops of the ribsections 14 directly opposed to each other via the air passage 18(theportions having the area S2) are arranged on a straight line in the airflowing direction A, a main air stream linearly passes a portion havinga reduced area S2 as shown by an arrow E.

Therefore, in the base plate section 13 constituting the enlargedportion having the passage area S1 in the heat transfer plate 12, aregion F in which air stream dwells is formed along the surface of thebase plate section 13, which significantly deteriorates the heattransfer rate on the surface of the base plate section 13.

In the second patent document, as the main stream of air moves straightpast the portion having the reduced area S2 as shown by an arrow E,unless the fin 17 is projected into the gap through which the mainstream E passes, it does not serve to improve the heat transferperformance.

Accordingly, in the second patent document, it is necessary that aheight of the fin 17 is higher than a height of a top of the rib section14, which forces the metallic material forming the heat transfer plate12 to be excessively stretched during the machining of the fin 17.Accordingly, it is difficult to machine the fins 17.

Also, if the fin 17 is made to be higher than the top of the rib section14, in a process of assembling the heat exchanger, the fin 17 is liableto collide with a peripheral member and be damaged.

SUMMARY OF THE INVENTION

In view of the above problems in the prior art, an object of the presentinvention is to improve the heat transfer performance of the plate typeheat exchanger in which separate fin members are not combined with heattransfer plates constituting internal passages, without increasing thenumber of heat transfer plates.

Also, another object of the present invention is both to improve theheat transfer performance of the plate type heat exchanger of this kindand to facilitate the production of the heat exchanger.

To achieve the above objects, according to the inventive heat exchanger,a plurality of heat transfer plates (12) forming plate surfacesextending in the flow direction (A) of an external fluid are stackedperpendicular to said plate surfaces,

a gap is provided between said plate surfaces of said adjacent heattransfer plates (12) to form an external passage (18) through which saidexternal fluid flows,

a plurality of rib sections (14) extending orthogonal to the flowingdirection A of said external fluid are projected from said platesurfaces into said external passage (18) to be integral with said heattransfer plates (12),

by shifting the positions of the plurality of rib sections (14) in oneof said adjacent heat transfer plates (12) relative to the positions ofthe plurality of rib sections (14) in the other of said adjacent heattransfer plates (12) as seen in the flowing direction A of said externalfluid, said external passage (18) is formed in a meandering manner,

the plurality of rib sections (14) form an internal passage (15, 16)inside thereof, through which flows an internal fluid,

fin sections (17) are projected from said plate surfaces at positionsbetween the adjacent rib sections (14) to be integral with the heattransfer plate (12), and

said fin section is press-formed (lances) so as to protrude a cutsurface partially cut a plate thickness of said heat transfer plate(12).

According to this structure, as the external fluid impinges on the ribsection (14) to generate turbulence, the local heat transfer rate isimproved in the vicinity of the rib section (14). Simultaneouslytherewith, the external passage (18) is formed in a meandering manner,whereby a main stream of the external fluid can positively impinge ontoa plate surface located between the plurality of rib sections (14).Thus, the local heat transfer rate is also improved on the plate surfacebetween the rib sections (14).

Further, due to the tip end effect (an effect of thinning thetemperature boundary layer) of the fin section (17), the local heattransfer rate of the fin section (17) is largely improved and anexternal fluid-side heat transfer area of the heat transfer plate (12)is increased by forming the fin section (17).

For the above-mentioned reasons, it is possible to improve the heattransfer rate of the plate type heat exchanger without increasing thenumber of heat transfer plates, and the practical merit thereof issignificant.

As a main stream of the external fluid impinges onto the plate surfacelocated between the plurality of rib sections (14) by forming theexternal passage (18) in a meandering manner, it is unnecessary to makethe fin section (17) higher than the rib section (14) as described inthe second patent document. Thus, it is possible to make the height ofthe fin section (17) lower than that disclosed in the second patentdocument.

Thus, when the fin section (17) is press-formed so as to protrude a cutportion partially cut the plate thickness of the heat transfer plate(12), the stretching of the heat transfer plate material becomes smallto allow easy formation of the fin section (17) and an inconvenience isavoided in that the fin section (17) impinges on the peripheral membersand is damaged during the assembly of the heat exchanger.

In the plate type heat exchanger which is the subject matter of thepresent invention, there is a problem of abnormal air noise (wind sound)caused by vortices generated downstream of the rib section (14) asdisclosed in Japanese Unexamined Patent Publication No. 2002-48491.According to the present invention, it is possible to shift (vary), inthe longitudinal direction of the rib section (14), the timing at whichthe external fluid stream moves over the rib section (14) by providingthe fin section (17) between the plurality of rib sections (14).

Thereby, as the overlap of sound waves based on the vortices generatedbehind the rib section (14) is suppressed to avoid the resonance, theabnormal air noise (wind sound) based on the provision of the ribsections (14) is reduced.

According to the present invention, said heat transfer plates (14) arecombined to form pairs, and said rib sections (14) and said fin sections(17) are formed integral with said pair of heat transfer plates (12),and the pair of heat transfer plates (12) are fixed together to formsaid internal passage (15, 16) inside the plurality of rib sections(14).

Thus, as the rib sections (14) and the fin sections (17) are formedintegral with the pair of heat transfer plates (12), the above-mentionedeffects are effectively exhibited.

Brazing is a representative means for fixing the pair of heat transferplates (12) to each other. When the fin sections (17) are provided, cutholes (17 d) are simultaneously formed, which operate as air-dischargingholes during the brazing, whereby the brazing between the pair of heattransfer plates (12) is improved.

In this regard, according to the present invention, the pair of heattransfer plates (12) includes two completely separated plates as well asa single plate folded at a center thereof to be two parts, each beinghalf the overall size.

According to the present invention, positions in the pair of heattransfer plates (12) at which said rib sections (14) are formed areshifted in the flowing direction (A) of the external fluid, and

said internal passage (15, 16) may be formed by said rib sections (14)formed in one of the pair of heat transfer plates (12) and a platesurface of the other.

According to the present invention, said rib sections (14) may be formedin said pair of heat transfer plates (12) at the same positions as seenin the flowing direction (A) of said external fluid, and

said internal passages (15, 16) are formed by the combination of saidrib sections (14) formed in said pair of heat transfer plates (12),respectively.

Thus, as the internal passages (15, 16) are formed by the combination ofthe rib sections (14) in the pair of heat transfer plates (12), it ispossible to increase the area of the internal passage in comparison withthe above-mentioned invention. Accordingly, it is possible to enlargethe mutual distance between the rib sections (14) and easily increasethe number of fin sections (1).

According to the present invention, said heat transfer plate (12) isconstituted by a single extrusion-formed plate material,

said rib sections (14) are formed by extrusion-forming a tubular-shapedportion on said single extrusion-formed plate material, and

said fin sections (17) are formed integral with said singleextrusion-formed plate material to be projected from a plate surface ofsaid single extrusion-formed plate material.

As the rib sections (14), that is the internal passages (15, 16) areformed by extrusion-forming a tubular-shaped portion on the singleextrusion-formed plate material, a coupling structure is unnecessary forthe purpose of forming the internal passages (15, 16). As a result, thecoupling portions in the heat exchanger as a whole are largely reducedto improve the productivity of the heat exchanger.

According to the present invention, said heat transfer plate (12) has abase plate section (13) having a flat surface between the adjacent ribsections (14), and

said fin section (17) is formed in said base plate section (13).

Thereby, the fin section (17) is easily formed on the flat surface ofthe base plate section (13).

According to the present invention, a width (Fw) in the flowingdirection (A) of said external fluid of said fin section (17) is 5 mm orless. Thereby, the tip end effect (effect for thinning the temperatureboundary layer) of the fin section (17) is effectively exhibited and theexternal fluid side heat transfer rate of the heat transfer plate (12)is advantageously improved.

According to the present invention, said fin section (17) is a slit finhaving an offset wall surface (17 a) set apart from a plate surface ofsaid heat transfer plate (12) with a predetermined gap, wherein saidoffset wall surface (17 a) are coupled to a plate surface of said heattransfer plate (12) at two positions.

By employing such a slit fin, it is possible to effectively improve theexternal fluid side heat transfer performance of the heat transfer plate(12).

When a gap, between positions on the pair of heat transfer plates (12)opposed to each other to define said external passage (18), at to whichpositions are formed said slit fins 17, is defined as (L1 to L3), and aprojected height of said offset wall surface (17 a) from a plate surfaceof said heat transfer plate (12) is defined as Fha1 to Fha3, thefollowing relationship is satisfied:Fha1 to Fha3≦½(L1 to L3).

If the offset wall surface (17 a) is formed thus, it is possible to forma meandering stream of the external fluid closer to a flat plate surfaceof the heat transfer plate (12). Thus, the impingement of the externalfluid to a surface of the heat transfer plate (12) is facilitated.

According to the present invention, a cross-sectional shape of said ribsection (14) has a curved surface, projected from the surface of saidheat transfer plate, which is generally semicircular,

said slit fin (17) is located at a position directly on downstream sidefrom said external fluid relative to said rib section (14), and

said offset wall surface (17 a) is inclined in the same direction as theinclination of the downstream side curved surface in the generallysemicircular curved surface of said rib section (14).

Thereby, it is possible to form a stream P approaching the downstreamside curved surface of the rib section (14) due to the guiding operationof the inclined surface of the offset wall surface (17 a) as illustratedin FIG. 25 described later. Thus, as the vortices (M′) are reduced tominimize the dwelling region caused by the vortices (M′), it is possibleto improve the heat transfer rate of the downstream side curved surfaceof the rib section (14) and of the flat surface of the heat transferplate (12).

According to the present invention, the cross-sectional shape of saidrib section (14) is such that it has a curved surface protrudedsemi-circularly from a surface of said heat transfer plate (12),

said slit fin is disposed adjacent to said rib section (14) at aposition directly on the upstream side of said external fluid, and

said offset wall surface (17 a) is inclined in the same direction as theinclination of the upstream side curved surface in a generallysemicircular curved surface of said rib section (14).

Thereby, as the inclined surface of the offset wall surface (17 a) isinclined in the same direction as the inclination of the upstream sidecurved surface of the rib section (14), it is possible to make theexternal fluid stream meander smoothly on the upstream side.

According to the present invention, said slit fin (17) is disposedopposite to a front of said rib section (14) while interposing saidexternal passage (18), and said offset wall surface (17 a) is formed tobe parallel to a flat surface of said heat transfer plate (12).

In this regard, as illustrated in FIG. 21 described later, a reversedposition of the stream in the external passage (18) is formed in frontof the rib section (14). If the offset wall surface (17 a) inclined in apredetermined direction is disposed at this reversed position of thestream, the inclination of the offset wall surface (17 a) prevents thereversing of the stream. However, in the present invention, as theoffset wall surface (17 a) is parallel to a flat surface of the heattransfer plate (12), the offset wall surface (17 a) becomes neutralrelative to the reversal of the stream and does not obstruct thereversal of the stream.

According to the present invention, said external fluid is air and saidinternal fluid is coolant for cooling said air, wherein said heatexchanger is constituted as a cooling heat exchanger generatingcondensation water on the surface of said heat transfer plate (12), and

a gap (Q1, Q2) between said offset wall surface (17 a) and the surfaceof said heat transfer plate (12) is 0.3 mm or more.

According to the study of the present inventors, it has been confirmedthat, when the gap (Q1, Q2) is 0.3 mm or more, the blockade of this gap(Q1, Q2) is avoidable and the drainage of condensation water isperformed.

According to the present invention, said fin section (17) is a protruded(lanced) fin to protrude at a predetermined angle relative to thesurface of said heat transfer plate.

The protruded (lanced) fin is simple in shape and easily formed incomparison with the slit fin defined by the above invention.

According to the present invention, said protruded fin (17) istriangular.

Such a triangular protruded fin (17), that is a delta wing, is liable togenerate a Karman vortex which improves the local heat transfer rate onthe periphery of the fin section due to the release of the Karmanvortex.

According to the present invention, said triangular protruded fin (17)is inclined to the flowing direction (A) of said external fluid at anangle from 15° to 45°.

According to the present invention, said protruded fin (17) isrectangular. Here, “rectangular” includes square and trapezoidal.

According to the present invention, the inclination angle of saidprotruded fin (17) relative to the flowing direction (A) of saidexternal fluid is determined in a small angle range from −30° to +30° sothat a surface of the protruded fin (17) follows the flowing direction(A) of said external fluid, whereby the ventilation resistance of theexternal fluid is reduced.

According to the present invention, said external fluid is air, andinternal fluid for cooling said air flows through said internal passage(15, 16),

said heat transfer plate (12) is disposed so that the longitudinaldirection of said rib section (14) coincides with the upward/downwarddirection, and

an inclination angle of said protruded fin (17) is in a range from 60°to 120° relative to the flowing direction (A) of said external fluid sothat a surface of said protruded fin (17) follows the longitudinaldirection of said rib section (14).

Thereby, when the condensation water generated on the surface of theheat transfer plate (12) drops in the longitudinal direction of theprotruded fin (17), the drainage of the condensation water isfacilitated since the protruded fin (17) does not disturb the droppingof the condensation water.

According to the present invention, said internal passage has anupstream side internal passage (16) disposed on the upstream side in theflowing direction (A) of said external fluid and a downstream sideinternal passage (15) disposed on the downstream side in the flowingdirection (A) of said external fluid,

said upstream side internal passage (16) and said downstream sideinternal passage (15) are respectively sectioned vertically to theflowing direction (A) of said external fluid into a plurality of areas(X, Y), and

passages connected in parallel to each other are constituted between theplurality of areas (X, Y) of said upstream side internal passages (16)and the plurality of areas (X, Y) of said downstream side internalpassages (15).

Thereby, it is possible to lower the pressure loss in the internalpassages (15, 16) as a whole by the parallel passage structure. Also, itis possible to reduce the number of rib sections (14) as well as enlargea gap between the heat transfer plates (12) laid to each other,resulting in the reduction of the external fluid side ventilationresistance.

According to the present invention, if said downstream side internalpassage (15) is an inlet side passage for said internal fluid, and saidupstream side internal passage (16) is an exit side passage for saidinternal fluid, a high efficiency orthogonally opposed type heatexchanger is obtained.

According to the present invention, if said parallel passages couple theplurality of areas (X, Y) in said upstream side internal passage (16)and the plurality of areas (X, Y) in said downstream side internalpassage (15) to each other in an X pattern, both the reduction ofpressure loss in the internal passage (15, 16) and the uniformity of thetemperature distribution of the blown-out external fluid are attainable.

Note, reference numerals in brackets indicate the correspondence of therespective members with concrete means in embodiments described later.

The present invention may be more fully understood from the descriptionof preferred embodiments of the present invention, as set forth below,together with the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

In the drawings:

FIG. 1 is an exploded perspective view of an evaporator according to afirst embodiment of the present invention;

FIG. 2 is an exploded perspective view illustrating a coolant flowpassage of the evaporator according to the first embodiment;

FIG. 3 is cross-section taken along a line III-III in FIG. 1;

FIG. 4 is a perspective view of part of a heat transfer plate shown inFIG. 3;

FIG. 5 is a perspective view of part of a core portion of a prior artfin and tube type heat exchanger;

FIG. 6 is a cross-section of a prior art finless type heat exchanger(shown in the first patent document);

FIG. 7 is a table showing the comparison of various items in the heatexchanger of the prior art with those in the first embodiment;

FIG. 8 is a graph of the local heat transfer rate in the finless typeheat exchanger shown in the prior art (the first patent document);

FIG. 9 is a cross-section of a core portion of an evaporator accordingto a second embodiment of the present invention;

FIG. 10A is a perspective view of part of a heat transfer plateaccording to a third embodiment of the present invention and

FIG. 10B is a perspective view of part of a heat transfer plateaccording to a comparative example for the third embodiment;

FIG. 11 is a cross-section of a core portion of an evaporator accordingto a fourth embodiment of the present invention;

FIG. 12 is a perspective view of part the heat transfer plate shown inFIG. 11;

FIG. 13 is an enlarged view of part of FIG. 12;

FIG. 14 is a perspective view of part of a heat transfer plate accordingto a fifth embodiment of the present invention;

FIG. 15 is an exploded perspective view illustrating a structure of acoolant flow passage according to a sixth embodiment of the presentinvention;

FIG. 16 is a schematic perspective view illustrating a structure of acoolant flow passage in an evaporator according to the sixth embodiment;

FIG. 17 is a perspective view of part of a heat transfer plateillustrating a fin shape according to a seventh embodiment of thepresent invention;

FIG. 18 is a cross-section of a core portion of an evaporator accordingto an eighth embodiment of the present invention;

FIG. 19 is a cross-section of part of a core portion of an evaporatoraccording to a ninth embodiment of the present invention;

FIG. 20 is a cross-section of part of a core portion of an evaporatoraccording to a tenth embodiment of the present invention;

FIG. 21 is a cross-section of part of a core portion of an evaporatoraccording to the tenth embodiment of the present invention;

FIG. 22 is a cross-section of part of a core portion of an evaporatoraccording to a comparative example of the tenth embodiment of thepresent invention;

FIG. 23 is a cross-section of part of a core portion of an evaporatoraccording to an eleventh embodiment of the present invention;

FIG. 24 is a cross-section of part of a core portion of an evaporatoraccording to a twelfth embodiment of the present invention;

FIG. 25A is an enlarged cross-section of part of a core portion of anevaporator according to a comparative example of the twelfth embodiment,and FIG. 25B is an enlarged cross-section of part of a core portionaccording to the twelfth embodiment;

FIG. 26 is a cross-section of part of a core portion of an evaporatoraccording to a thirteenth embodiment of the present invention; and

FIG. 27 is a cross-section of a main part of a heat exchanger accordingto a second patent document.

DESCRIPTION OF THE PREFERRED EMBODIMENTS First Embodiment

A first embodiment is an evaporator for a vehicle air conditioner.Initially, the total structure of the evaporator 10 for a vehicle airconditioner will be described. FIG. 1 is an exploded perspective viewillustrating a summary of the total structure of the evaporator, andFIG. 2 is an exploded perspective view wherein a coolant passageindicated by arrows is added to FIG. 1. FIG. 3 is a lateralcross-section illustrating a laminated structure of heat transfer plates12, and is a cross-section taken along a line I-I in FIG. 1. FIG. 4 isan enlarged perspective view of part of a heat transfer plate 12.

The total structure of the evaporator shown in FIGS. 1 and 2 may besubstantially the same as that disclosed in the above-mentioned firstpatent document (Japanese Unexamined Patent Publication No. 11-287580).The evaporator 10 is an orthogonal opposed-flow-type heat exchangerwherein a flowing direction A of conditioning air and a flowingdirection B of coolant in a heat transfer plate (up-down direction inFIG. 1) are orthogonal to each other, and the upstream (inlet) sidepassage of the coolant stream is located downstream from the air flowingdirection A, while the downstream (exit) side passage of the coolant islocated upstream from the air flowing direction A. In this regard, inthe evaporator 10, air is an external fluid (cooled fluid) and coolantis an internal cooling fluid.

This evaporator 10 constitutes a core section 11 for carrying out theheat exchange between the conditioned air and the coolant solely bystacking a number of heat transfer plates 12 in the direction verticalto the plate surface (in the direction orthogonal to the air flowingdirection A). In this regard, at the uppermost and lowermost ends ofthese heat transfer plates 2, tanks 20 to 23, described later, areformed. As a portion in which the tank 20 to 23 is formed does not allowair to pass therethrough, the core section 11 is formed in anintermediate region of the heat transfer plate 12 except for the tanks20 to 23 formed at the upper and lower ends.

The respective heat transfer plate 12 is press-formed from a metallicsheet and, more concretely, from A3000-type aluminum core material cladwith A400-type aluminum material on both side surface thereof. A platethickness t of the heat transfer plate 12 (FIG. 2) is as small as, forexample, 0.15 mm. The heat transfer plates 12 have a generallyrectangular planar shape having the same dimensions.

Next, a concrete shape of the heat transfer plate 12 will be describedwith reference to FIG. 3. The respective heat transfer plate 12 has ribsections 14 formed from a flat base plate 13 by the press-forming. Therib sections 14 are of a longitudinal shape continuously extending inparallel to each other in the longitudinal direction of the heattransfer plate 12. While a cross-sectional shape of the rib section isgenerally semicircular in FIG. 3, it may be other shapes, for example, atrapezoid having rounded corners.

An inside space of the rib section 14 forms an internal passage, moreconcretely, a coolant passage 15, 16, through which flows a low pressureside coolant after passing through a pressure-reduction means (anexpansion valve or others) in a refrigerant cycle. As the longitudinaldirection of the heat transfer plate 12 coincides with upward anddownward, the longitudinal direction of the rib sections 14 is alsocoincides with upward and downward; that is, orthogonal to the airflowing direction A.

At a central position of a rib pitch Rp which is a mutual distancebetween the adjacent rib sections 14 in one heat transfer plate 12,there the rib section 14 in the other heat transfer plate 12 matedtherewith. Accordingly, when the pair of heat transfer plates 12, 12 arelocated so that the rib sections 14 of the respective heat transferplates are opposed to each other toward outside and the base platesections 13 are in contact with each other, the inner side of the ribsection in the one heat transfer plate 12 is tightly closed by a centralwall surface of the base plate section 13 in the other heat transferplate 12.

Accordingly, coolant passages 15 and 16 are formed between therespective inner side of the rib section 14 and the base plate section13 in the mated heat transfer plate 12. The coolant passage 15constitutes a leeward coolant passage disposed on the downstream sidearea of the air flowing direction A, while the coolant passage 16constitutes a windward coolant passage disposed on the upstream sidearea of the air flowing direction A.

A fin section 17 is formed integrally at a position at which the baseplate sections 13 in the respective heat transfer plates 12, 12 are incontact with each other. The fin section 17 is formed between theadjacent rib sections 14. In this embodiment, the fin sections 17 in thepair of heat transfer plates 12, 12 are formed at the same position asseen in the air flowing direction A.

The fin section 17 in this embodiment constitutes a slit fin. The slitfin is one having an offset wall surface 17 a which is a top wallsurface apart from a surface of a mother material (concretely, a surfaceof the base plate section 13), and at a predetermined gap, to define aspace for allowing air to pass therethrough between the offset wallsurface 17 a and the surface of the mother material as shown in FIG. 4,wherein at least two positions of the offset wall surface 17 a arephysically fixed to the surface of the mother material.

In the embodiment shown in FIG. 4, the fin section 17 is of a U-shape inwhich left and right ends of the offset wall surface 17 a are fixed tothe base plate section 13 with two lateral walls 17 b and 17 c.

In this regard, a fin height Fh which is a height of the offset wallsurface 17 a of the fin section 17 is the same as a height (rib height)Rh of the rib section 14 or slightly lower than height Rh as shown inFIG. 3. In the embodiment shown in FIG. 4, the slit fin 17 has awidthwise dimension Fw in the air flowing direction A smaller than adimension orthogonal to the air flowing direction A (an upward/downwarddimension in FIG. 4).

To form such a fin section 17, two cut lines are provided in afin-forming area of the base plate section 13 at a distancecorresponding to the fin width Fw, after which a region between the twocut lines is pressed to have a U-shaped cross-section.

The U-shape (a slit fin shape) of the fin section 17 constitutes aprotruded shape having a cut surface passing through a plate thicknessof the heat transfer plate 12. Thereby, a cut hole 17 d accompanied withthe formation of the fin section 17 is formed in the fin-forming area ofthe base plate section 13.

In this regard, as the fin-forming area is provided at a position atwhich the base plate sections 13 in the pair of heat transfer plates 12,12 are in contact with each other, there is no risk in that the coolantleaks from the coolant passages 15 and 16 even if the cut hole 17 d isformed in the base plate section 13.

In this regard, in FIGS. 1 and 2, the above-mentioned fin sections 17are not shown for simplifying the illustration. In FIG. 3, the number ofthe rib sections 14 in the pair of heat transfer plates 12, 12 is five.On the other hand, in FIGS. 1 and 2, the number of the rib sections 14in one of the pair of heat transfer plates 12, 12 is six, and that inthe other of the pair is five. The number of the rib sections 14, thatis, the number of coolant passages 15, 16 may, of course, be increasedor decreased in accordance with the required performance or contour ofthe evaporator 10.

At each of opposite end areas of the respective heat transfer plate 12as seen in the direction B (the longitudinal direction of the heattransfer plate) orthogonal to the air flowing direction A, two tanksections 20 to 23 divided in the widthwise direction of the heattransfer plate (in the air flowing direction A) are formed. That is,there are two tank sections 20 and 22 at the upper end area of the heattransfer plate 12, and two tank sections 21 and 23 at the lower end areathereof.

The tank sections 20 to 23 are formed to be projected in the samedirection as the rib sections 14 in the respective heat transfer plate12. A projected height of the tank section 20 to 23 is one half of atube pitch Tp (see FIG. 3) so that tops of the adjacent tank sections 20to 23 are brought into contact and fixed with each other.

In this regard, the projected height of the tank section 20 to 23includes the plate thickness t of the heat transfer plate 12. The tubepitch Tp is a distance between the arranged heat transfer plates 12.Also, a space pitch Sp is a value obtained by subtracting the platethicknesses t of two heat transfer plates 12 from the tube pitch Tp;i.e., Tp−2t.

In the embodiment shown in FIG. 3, while the rib height of the ribsection Rh is determined to be one half of the tube pitch Tp, that is,generally equal to the projected height of the respective tank section20 to 23, this is not limitative, but the rib height Rh of the ribsection 14 may be slightly increased or decreased relative to therespective tank section 20 to 23.

As mentioned above, as the tank sections 20 to 23 are projected in thesame direction as the rib sections 14, and recessed longitudinalopposite end areas formed by the projection of the rib sections 14 arecontinuous to the recessed shape of the tank sections 20 to 23, both endportions of the windward coolant passage 16 communicate with the leewardupper and lower tank sections 22, 23, and both ends portions of theleeward coolant passage 15 communicate with the windward upper and lowertank sections 20, 21.

In this regard, the leeward tank section 20 and the windward tanksection 22 on the upper side of the heat transfer plate define thecoolant passages independent from each other, and the leeward tanksection 21 and the windward tank section 23 on the lower side of theheat transfer plate define the coolant passages independent from eachother.

As communication openings 20 a to 23 a are provided at the centers ofthe tops of the respective tank sections 20 to 23, it is possible tocommunicate the communication openings 20 a to 23 a with each other bybringing the projected tops of the tank sections 20 to 23 adjacent toeach other and fixing them together.

Thereby, it is possible to mutually communicate the coolant passages ofthe tank sections 20 to 23, between adjacent heat transfer plates, asseen in the left/right direction in FIGS. 1 and 2.

As the plurality of rib sections 13 in the respective heat transferplate 12 are disposed while being shifted from those in the adjacentheat transfer plate 14 as seen in the widthwise direction of the heattransfer plate 12 (in the air flowing direction A) as shown in FIG. 3,it is possible to oppose the respective rib section 14 to the base platesection 13 in the adjacent heat transfer plate 12.

As the rib height Rh of the rib section 14 is determined to beapproximately equal to half a tube pitch Tp as described before, a gapis formed between a top of the rib section 14 on the convex side and thebase plate section 13 in the adjacent heat transfer plate 12, whereby ameandering curved air passage 18 is continuously formed along a totallength (in the air flowing direction A) of the heat transfer plate 12 asshown in FIG. 3 by an arrow A1. The fin sections 17 constitutingU-shaped slit fins are arranged in this wavy air passage 18 adjacent tothe respective rib sections 14.

Then, a portion for feeding and discharging coolant relative to the coresection 11 is described below. As shown in FIGS. 1 and 2, end plates 24,25 having the same size as the heat transfer plate 12 are disposed atopposite ends in the laying direction of the heat transfer plates. Theend plate 24, 25 is a flat plate capable of being in contact with aconvex side of the tank sections 20 to 23 of the heat transfer plate 12and fixed thereto.

Into holes provided in the vicinity of the upper end of the left sideend plate 24 in FIGS. 1 and 2, a coolant inlet pipe 24 a and a coolantexit pipe 24 b are fixed, wherein the coolant inlet pipe 24 acommunicates with a communication opening 20 a formed at a top of theleeward side tank section 20 formed at an upper end of the leftmost heattransfer plate 12. The coolant exit pipe 24 b communicates with acommunication opening 22 a formed at a top of the windward tank section22 formed at an upper end of the leftmost heat transfer plate 12.

The left side end plate 24 is formed of a both-side aluminum cladmaterial in the same manner as in the heat transfer plate 12, and brazedto the coolant inlet and exit pipes 24 a, 24 b. The right side end plate25 is formed of metallic material clad with brazing metal on one side tobe brazed to the heat transfer plate 12.

A liquid-vapor-type two-phase coolant passing through apressure-reduction means such as an expansion valve is fed into thecoolant inlet pipe 24 a. On the other hand, the coolant exit pipe 24 bis connected to a suction side of a compressor, not shown, so thatevaporated vapor-liquid type coolant evaporated in the evaporator 10 isguided to the suction side of the compressor.

In a group of the plurality of heat transfer plates 12 stacked to eachother in the left/right direction in FIGS. 1 and 2, the leeward sidecoolant passage 15 formed in the interior of the rib sections 14described before constitutes the inlet side coolant passage in all ofthe evaporator as the coolant is fed from the coolant inlet pipe 23.

On the other hand, the windward side coolant passage 16 formed in theinterior of the rib sections 14, described before, constitutes theoutlet side coolant passage directing the coolant passing through theleeward side (inlet side) coolant passage 15 as the coolant is fed tothe coolant outlet pipe 24 b.

Next, all the coolant passages in the evaporator 10 will be describedwith reference to FIG. 2. The leeward tank sections 20 and 21 in thetank sections 20 to 23 disposed at the upper and lower ends of theevaporator 10 constitute the coolant inlet side tank sections, while thewindward tank sections 22 and 23 constitute the coolant exit side tanksections.

The leeward and upper side coolant inlet side tank section 20 is dividedby a partition (not shown) disposed at an intermediate position in thelaying direction of the heat transfer plates 12 into the left side flowpassage in FIG. 2 (a flow passage on the area X side) and the right sideflow passage in FIG. 2 (a flow passage on the area Y side).

Similarly, the windward and upper side coolant exit side tank section 22on is divide into the left side flow passage in FIG. 2 (a flow passageon the area X side) and the right side flow passage in FIG. 2 (a flowpassage on the area Y side). These divided portions are simplyconstituted by using the heat transfer plate 12 among those describedbefore, which are located at the intermediate position and thecommunication openings at the tops of the tank sections 20 and 22thereof are blocked to be a barrier wall (a blind lid).

According to the coolant passage structure in FIG. 2, the vapor-liquidtype two phase coolant depressurized by the expansion valve enters theleeward upper side inlet side tank section 20 from the coolant inletpipe 24 a as shown by an arrow a. As the flow passage of the inlet sidetank section 20 is divided into the left and right areas X and Y by thepartition not shown, the coolant is fed solely into the left side area Xof the inlet side tank section 20.

In the left side area X in FIG. 2, the coolant descends in the coolantpassage 15 formed by the leeward side rib sections 14 as indicated by anarrow b and enters the lower side inlet side tank section 21. Next, thecoolant moves through the lower side inlet side tank section 21 into theright side area Y in FIG. 2 as indicated by an arrow c, and rises in thecoolant passage 15 formed in the right side area Y by the leeward siderib section 14 of the heat transfer plate 12, as indicated by an arrow dto enter the right side area Y of the upper side inlet side tank section20.

Here, the communication opening 20 a of the inlet side tank section 20in the rightmost side heat transfer plate 12 communicates with thecommunication opening 22 a of the exit side tank section 22 located onthe upper side of the rightmost side heat transfer plate 12 via acommunication passage (not shown, see an arrow f) formed in the vicinityof the upper end of the right side end plate 25.

Accordingly, the coolant entering the flow passage of the right sidearea Y in the upper side inlet side tank section 20 flows rightward asindicated by an arrow e, and thereafter, passes the communicationpassage (not shown) in the vicinity of the upper end of the right sideend plate 25 as indicated by an arrow f and enters the flow passage inthe right side area Y of the upper side exit side tank section 22.

Here, as the flow passage of the exit side tank section 22 is dividedinto the left and right side areas X and Y by the above-mentionedpartition not shown, the coolant solely enters the flow passage in theright side area Y of the exit side tank section 22 as indicated by anarrow g. Next, the coolant entering the right side area Y in this tanksection 22 descends the coolant passage 16 formed by the windward ribsection 14 in the heat transfer plate 12 and enters the right side areaY of the lower side exit side tank section 23.

The coolant moves from the right side area Y to the left side area X inFIG. 2 through the lower side exit side tank section 23 as indicated byan arrow i, and thereafter, rises the coolant passage 16 formed by thewindward rib section 14 of the heat transfer plate 12 as indicated by anarrow j to enter the flow passage in the left side area X of the upperside exit side tank section 22. The coolant passes through the exit sidetank section 22 leftward as indicated by an arrow k, and is dischargedfrom the coolant exit pipe 24 b to the outside of the evaporator.

In the evaporator 10 shown in FIGS. 1 and 2, the coolant passage isstructured as described above, whereby it is possible to assemble theevaporator 10 by laying the respective components (12, 24, 25, 24 a and24 b) to be in contact with each other, holding such a stacked state(assembled state) by a suitable jig, putting the assembly into thebrazing furnace, and heating the same to a melting point of the brazingmaterial. Thus, the assembly of the evaporator 10 is completed.

Next, the operation of the above-mentioned evaporator 10 will bedescribed. The evaporator 10 shown in FIGS. 1 and 2 is accommodated in acase of an air conditioning unit not shown, upside down, so that airflows in the direction A due to the action of an air conditioningblower.

When the compressor for the refrigeration cycle operates, the lowpressure vapor-liquid type two-phase coolant decompressed by theexpansion valve not shown flows via the above-mentioned passagesindicated by the arrows a to k shown in FIG. 2. On the other hand, dueto the gap formed between the rib section 14 projected on the outersurface side of the heat transfer plate 12 and the base plate section13, the air passage 18 meandering as shown by an arrow A1 in FIG. 3 iscontinuously formed in the widthwise direction of the heat transferplate (the air flowing direction A).

As a result, the conditioned air sent in the direction A passes throughthe air passage 18 between the two heat transfer plates 12 and 12 whilemeandering as shown by the arrow Al. Since the coolant sucks theevaporation latent heat from this air stream and vaporizes, theconditioned air fed in the direction A becomes a cold wind.

At this time, as the inlet side coolant passage 15 is arranged on theleeward side and the exit side coolant passage 16 is arranged on thewindward side relative to the flowing direction A of the conditionedair, the relationship of the coolant inlet and exit relative to the airflow becomes a counter flow relationship.

Further, as the air flowing direction A is orthogonal to thelongitudinal direction of the rib section 14 of the heat transfer plate12 (the coolant flowing direction B in the coolant passage 15, 16) onthe air side, and the rib section 14 forms a convex heat transfersurface projected orthogonal to the air flow, the straight advancementof air is disturbed by this orthogonally extended rib section 14.Accordingly, the air stream is disturbed to be turbulent, whereby theair-side heat transfer rate is significantly improved.

In this regard, in the plate type heat exchanger wherein the coresection 11 is structured solely by the heat transfer plates 12 as inthis embodiment, there is a problem in that the air-side heat transferarea is largely reduced in comparison with the conventional fin-and-tubetype heat exchanger, whereby it is difficult to ensure the required heattransfer performance.

In view of such a point, the present inventors have studied variouscountermeasures. For example, it is thought that the air-side heattransfer rate is improved by increasing a rib height Rh of the ribsection 14 to further facilitate the generation of turbulence. However,as the ventilation resistance becomes naturally larger if the rib heightRh increases, it is impossible to improve the performance in view of theventilation resistance ratio. Similarly, as the increase in the numberof the rib sections 14 results in the large ventilation resistance, itis impossible to improve the performance in view of the ventilationresistance ratio.

Although the reduction of the tube pitch Tp is advantageous forimproving the heat transfer performance, this is defective in that thenumber of heat transfer plates 12 increases, resulting in a largerweight of heat exchanger as well as a larger ventilation resistance.

Under such circumstances, according to this embodiment, a fin section 17constituting a U-shaped slit fin is provided at a position between theevery adjacent rib sections 14, that is, a position corresponding to theflat base plate section 13.

According to this structure, as air flows along inner and outer surfacesof the U-shaped fin section 17 whereby the inner and outer surfaces ofthe U-shaped fin section 17 become the air-side heat transfer area, theair-side heat transfer area is largely increased in comparison with onehaving no fin section 17.

In addition thereto, it is possible to effectively improve the air-sideheat transfer rate of the heat transfer plate 12. That is, while theair-side heat transfer rate is liable to be reduced in the base platesection 13 as a temperature boundary layer grows to be thick on a flatsurface of the base plate section 13 in the heat transfer plate 2 in theair flowing direction A, it is possible to divide the temperatureboundary layer on the flat surface of the base plate section 13 torestrict the growth of the temperature boundary layer by providing thefin section 17. Also, the air-side heat transfer rate of the fin section17 itself is sufficiently improved by the tip-end effect of the finsection.

Further, due to the meandering of the air stream in the air passage 18as shown in FIG. 3 by an arrow Al, it is possible to alternately impingea main air stream on the surface of the rib section 14 and the flatsurface of the base plate section 13. Thereby, it is possible to improvethe air-side heat transfer rate in the base plate section 13 by thinningthe temperature boundary layer on the flat surface of the base platesection 13.

As described above, according to this embodiment, it is possible toeffectively improve the heat transfer performance of the plate type heatexchanger while restricting the increase of the ventilation resitance.

The improvement effect of the heat transfer performance according tothis embodiment will be concretely described below in comparison withthe conventional fin-and-tube type heat exchanger and the finless typeheat exchanger disclosed in the first patent document.

FIG. 5 is a perspective view of part of a core portion of theconventional fin and tube type heat exchanger, wherein a corrugated fin51 is fixed between flat tubes 50A and 50B FIG. 5.

FIG. 6 is a cross-section of the finless type heat exchanger shown inthe first patent document, corresponding to a cross-section taken alonga line I-I in FIG. 1, having no fin section 17 according to thisembodiment.

FIG. 7 is a table illustrating the comparison of various items of theconventional fin-and-tube type heat exchanger (1) shown in FIG. 5, thefinless type heat exchanger (2) according to the first patent documentshown in FIG. 6 and the inventive heat exchanger according to thisembodiment shown in FIGS. 3 and 4. In this Table, each of the itemvalues of the fin-and-tube type heat exchanger (1) is selected as areference value (100), and the item values of the heat exchangers (2)and (3) are represented as ratios to the reference values of the heatexchanger (1).

The items in FIG. 7 are calculated in accordance with the followingconditions.

Contour size of heat exchanger: Width W 260 mm×height H 215 mm×depth D38 mm

Note the width W is a dimension in the plate-stacking direction and thedepth D is a thickness dimension in the air flowing direction.

Air flow: 500 m³/h, the ventilation resistance in the core section isthe same in the heat exchangers (1) to (3).

Fin pitch fp: 2.6 mm and fin height fh: 6 mm in the heat exchanger (1).

Thickness t: 0.15 mm, space pitch Sp: 2.6 mm, pitch Rp of the ribsections: 7.1 mm and height Rh of the rib section: 1.45 mm in the heattransfer plate 12 in the heat exchanger (2).

Thickness t: 0.15 mm, space pitch Sp: 3.0 mm, pitch Rp of the ribsections: 7.1 mm, height Rh of the rib section: 1.45 mm: fin height Fh:1.0 mm, fin width Fw: 0.8 mm in the heat transfer plate 12 in the heatexchanger (3). Note the fin pitch Fp is one half of the pitch Rp.

As is apparent from a comparison of the items in the respective heatexchangers (1), (2) and (3) shown in FIG. 7, in the finless type heatexchanger (2) disclosed in the first patent document, while the air-sideheat transfer rate is largely improved relative to the fin-and-tube typeheat exchanger (1), there is a defect in that the air-side heat transferarea Fa is largely reduced.

FIG. 8 illustrates the variation of the air-side local heat transferrate in the finless type heat exchanger (2) of the first patentdocument. As the air stream impinges on a convex front surface of therib section 14 in the heat transfer plate 12 to become turbulent, thelocal heat transfer rate is largely improved. On the contrary, on theflat surface of the base plate section 13 having no rib section 14, itis apparent that the temperature boundary layer grows to largelydeteriorate the local heat transfer rate.

Contrarily, according to this embodiment, the fin section 17constituting the U-shaped slit fin is disposed at a position between theadjacent rib sections 14 in the heat transfer plate 12; i.e., in theflat base plate section 13. As the heat transfer area of the heattransfer plate 12 is significantly increased by the formation of the finsection 17 and the temperature boundary layer on the flat surface of thebase plate section 13 is divided by the fin section 17 and thinned dueto the tip end effect, the heat transfer rate in the base plate section13 is also improved.

As mentioned above, according to the heat exchanger (3) of thisembodiment, it is possible to largely increase the air-side heattransfer area Fa and, simultaneously therewith, to improve the air-sideheat transfer rate αa, in comparison with the finless type heatexchanger (2) disclosed in the first patent document, as illustrated inFIG. 7.

In this regard, in the heat exchanger (3) of this embodiment, the finsection 17 is added to the heat exchanger (2). Therefore, if thestructure is as it is, the ventilation resistance increases. Actually,the space pitch Sp is increased from 2.6 mm in the heat exchanger (2) to3.0 mm. Accordingly, it is possible to equalize the ventilationresistance of the inventive heat exchanger (3) to that in the heatexchanger (2) as described in the preconditions for the calculationdescribed before.

According to the heat exchanger (3) in this embodiment, the number ofheat transfer plates 12 to be used is reduced due to the enlargement ofthe space pitch Sp. Thereby, an area of the coolant passage becomessmaller than that in the heat exchanger (2), and the in-tube side heattransfer rate αr is more improved than in the heat exchanger (2).

When the heat exchanger is structured as an evaporator 10 for coolingair as in this embodiment, the moisture in air is condensed due to thecooling action of the evaporator 10 to generate condensation water. Thedrainage of this water is an important problem.

In the finless type heat exchanger (2) and the inventive heat exchanger(3) in this embodiment, the air stream impinges onto a front side of theconvex of the rib section 14 extending upward/downward and generates thecondensation water which moves to a rear side of the convex of the ribsection 14 due to a wind pressure of the air stream and drops down alongthe rear side of the convex surface of the rib section 14 due to thegravity.

At that time, as fin section 17 is disposed more behind the convexsurface of the rib section 14 in the heat exchanger (3) of thisembodiment, condensation water is favorably drained along the rear sideof the convex surface of the rib section 14 even if the fin section 17is provided. Thus, an inconvenience, such as the increase in ventilationresistance caused by the dwell of the condensation water within the coresection, is avoidable.

Further, according to this embodiment, an effect is obtainable in thatabnormal air noise (wind sound) generated behind the rib section 14(downstream of the air flow) is minimized due to the existence of thefin section 17.

That is, as described in Japanese Unexamined Patent Publication No.2002-48491, in the finless type heat exchanger (2) disclosed in thefirst patent document, a layer peeled off from the main air stream isgenerated at the rear end of the rib section 14 as seen in the airflowing direction, and generates vortices therein. Further, as the ribsections 14 linearly extend orthogonal to the air flowing direction Awhile maintaining the same height, they generate vortices at the sametime at the rear ends of the rib sections 14. The simultaneousgeneration of vortices coupled in the longitudinal direction of the ribsections causes the overlap of sound waves to amplify the abnormal airnoise (wind sound).

On the contrary, according to this embodiment, as a position at whichthe fin section 17 is formed and a position at which the fin section isnot formed are alternately present in the widthwise direction of the airpassage 18 (in the longitudinal direction of the rib section 14), thevariation occurs in the air stream in the longitudinal direction of therib section 14, whereby it is possible to shift the timing at which theair stream moves over the rib section 14 in the longitudinal directionof the rib section 14. Thus, it is possible to divide vortices generatedat the rear end of the rib section 14 as seen in the air flowingdirection.

Thereby, as the overlap of sound waves based on the vortices generatedat the rear ends of the rib sections 14 is restricted to avoid theresonance, the abnormal air noise (wind sound) caused by the ribsections 14 is suppressed.

As a result, it is possible to restrict the generation of the overlap ofsound waves in the longitudinal direction of the rib section 14 andsuppress the resonance phenomenon, resulting in the reduction ofabnormal air noise (wind sound).

Also, according to this embodiment, as the fin sections 17 are pressedinto a shape from the base plate section 13 of the heat transfer plate12, the a hole 17 d is formed in the base plate section 13 at a positionat which the fin section 17 is formed. Due to this cut hole 17 d, anadditional effect is obtained to improve the brazing function of theheat exchanger.

That is, as shown in FIG. 3, according to this embodiment, a relativelywide fixing surface is formed between the adjacent rib sections 14wherein flat surfaces of the base plate sections 13 in two heat transferplates 12 are brought into contact with each other. In such a relativelywide fixing surface, the brazing defect is liable to occur due to theexistence of air layers in micro-gaps of the fixing surface. In thisembodiment, however, as the cut holes 17 d operate as air-dischargingholes for discharging air on the fixing surface, the base plate sections13 are favorably brazed to each other via the relatively wide fixingsurface.

According to the above-mentioned first embodiment, the basicconfiguration of the heat transfer plate 12 is a plain plate disposed toform a flat surface in the air flowing direction A, and the rib section14, the fin section 17 and the tank section 20 to 23 are formed in thisplain plate. An intermediate portion of the heat transfer plate 12,except for the upper and lower end tank sections 20 to 23; that is, thecore section 11, may not be a flat surface but may be a wavy surface (acurved surface gradually meandering in a wavy form). Even in such astructure, the same operation and effect are obtainable as in the firstembodiment.

Second Embodiment

The rib sections 14 in the two heat transfer plates 12 fixing the baseplate sections 13 to each other are disposed at positions shifted fromeach other in the air flowing direction A in the first embodiment.Contrarily, in a second embodiment, as shown in FIG. 9, the rib sections14 in the two heat transfer plates 12 fixing the base plate sections 13to each other are disposed at the same position in the air flowingdirection A.

In the second embodiment, as the rib sections 14 having a semicircularcross-section in the two heat transfer plates 12 are combined at thesame position to form circular coolant passages 15 and 16, the passagearea of the respective coolant passage 15, 16 becomes larger.

Thereby, it is possible to decrease the number of the rib portions 14 tolengthen a mutual distance between the adjacent rib sections 14; i.e. alength of the base plate section 13 in the air flowing direction.Therefore, as shown in FIG. 11, it is possible to arrange three finsections 17 between the adjacent rib sections 14.

According to the second embodiment, in accordance with the increase inthe passage areas of the coolant passages 15 and 16, the coolant flowspeed becomes lower and, as a result, the in-tube side heat transferrate αr is smaller than that in the first embodiment. However, as theair side heat transfer performance is improved due to the increase ofthe number of the fin sections 17, and compensates for the reduction ofthe coolant side heat transfer performance, the heat transferperformance becomes better as a whole than in the first embodiment.

In this regard, it is, of course, possible to variously increase ordecrease the number of the fin sections 17 in accordance with thespecifications of the evaporator 10.

Third Embodiment

In the second embodiment, the rib sections 14 having a semicircularcross-section in the two heat transfer plates 12 are combined togetherat the same position to obtain the coolant passages 15 and 16 having acircular cross-section. According to a third embodiment, as shown inFIG. 10A, tubular coolant passages 15 and 16 having a circularcross-section are formed in a single heat transfer plate 12 byextrusion. Due to this tubular shape, rib sections 14 having asemicircular cross-section are projected from front and rear surfaces ofthe single heat transfer plate 12.

After this extrusion process, the fin sections 17 are pressed from aflat surface of the base plate section 13 between the adjacent ribsections 14. In the embodiment shown in FIG. 10A, the fin section 17 isformed as a U-shaped slit fin.

According to the third embodiment, as the tubular coolant passages 15and 16 are formed in a single heat transfer plate 12 by the extrusion,the number of heat transfer plates 12 to be stacked is halved. Thereby,the positions necessary for brazing are largely decreased to improve theproductivity of the heat exchanger to a large extent.

FIG. 10B illustrates a comparative example of the third embodiment,wherein the fin sections 17 are not formed. As the front and rearsurfaces of a single heat transfer plate 12 are the air side heattransfer surface in this comparative example together with the thirdembodiment, even if the fin sections 17 are pressed as in the thirdembodiment, a large increase in the air side heat transfer area is notexpected.

However, as the heat transfer rate in the base plate section 13 islargely improved by the tip end effect derived from the provision of thefin sections 17, it is possible to realize an improvement in the heattransfer performance as a whole.

Fourth Embodiment

In the first to third embodiments, while the description has been madeon the structure wherein the fin section 17 is a U-shaped slit finhaving an offset wall surface 17 a, the fin section 17 should not belimited to the slit fin but may be a simple protruded fin. Here, theprotruded fin is one which is coupled to a mother material surface(concretely the surface of the base plate section 13) at at least onepoint and protruded by pressing to have a predetermined angle to themother material surface.

In the fourth embodiment, as shown in FIGS. 11 and 12, the fin section17 is a triangular fin protruded of a triangular piece cut at a rightangle from the flat surface of the base plate section 13. Due to thisprotruding of the triangular fin, a cut hole 17 d is formed on the flatsurface of the base plate section 13. This cut hole 17 d serves fordischarging air when the brazing is carried out.

The fin sections 17 are provided at the same position in the two heattransfer plates 15 and 16 constituting the coolant passages 15 and 16(the same position in the air flowing direction A). Also, the triangularpiece constituting the fin section 17 is inclined at a predeterminedangle 0 relative to the air flowing direction A. FIG. 13 is an enlargedview illustrating such a slanted arrangement of fin section 17.

The triangular fin section 17 constitutes a delta wing which is liableto generate a Karman vortex. In this regard, if the inclination angle θof the fin section 17 constituting the delta wing is determined in arange from 15 to 45°, it is possible to facilitate the effect forimproving the heat transfer rate in the base plate section 13 by thegeneration of the Karman vortex.

While the projected height of the fin section 17 is larger than half atube pitch Tp in FIG. 11, the projected height may, of course, beincreased or decreased if necessary and to, for example, smaller thanhalf a tube pitch Tp.

The protruded fin (the fin section 17) according to the fourthembodiment is not limited to a triangle but may be other shapes, such asa rectangle.

If the protruded fin (the fin section 17) of the fourth embodiment isarranged generally parallel to the air flowing direction A, it isadvantageous for reducing the ventilation resistance. Here, “generallyparallel to the air flowing direction A” means that the inclinationangle θ is within a range from −30 to +30°.

As a surface of the protruded fin (the fin section 17) becomes generallyparallel to the longitudinal direction of the rib section 14 (that is,upward/downward direction of the evaporator) if the protruded fin (thefin section 17) of the fourth embodiment is arranged generallyorthogonal to the air flowing direction A, the discharge of thecondensation water is hardly disturbed by the protruded fin (the finsection 17) when the condensation water drops down in the longitudinaldirection of the rib of the rib section 14. Here, “generally parallel tothe longitudinal direction of the rib section 14” means that theinclination angle θ is in a range from 60 to 120°.

Fifth Embodiment

While a plurality of fin sections 17 constituted by slit fins arelinearly arranged parallel to the air flowing direction A in the firstembodiment, a plurality of fin sections 17 constituted by slit fins arearranged in a zigzag manner relative to the air flowing direction A asshown in FIG. 14 according to a fifth embodiment. Here, the zigzagarrangement means that the plurality of fin sections 17 are arrangedwhile being shifted with respect to each other in the directionorthogonal to the air flowing direction A.

In this regard, when the fin section 17 is constituted by the protrudedfin as in the fifth embodiment, the fin sections 17 may be arranged in azigzag manner.

Sixth Embodiment

According to the first embodiment, coolant passages indicated by arrowsa to k are arranged in series between the coolant inlet pipe 24 a andthe coolant exit pipe 24 b as shown in FIG. 2. Contrarily, in a sixthembodiment, two coolant passages are arranged in parallel between thecoolant inlet pipe 24 a and the coolant exit pipe 24 b.

The sixth embodiment will be explained with reference to FIGS. 15 and16, wherein FIG. 15 is an exploded perspective view corresponding toFIG. 2, and FIG. 16 is a schematic perspective view illustrating thecoolant passages in the FIG. 15.

According to the sixth embodiment, a tank section 20 located on thedownstream side of the air flow and a tank section 22 located on theupstream side of the air flow are formed at an upper end of the heattransfer plate 12 as in the first embodiment. Contrarily, at a lower endof the heat transfer plate 12, a tank section divided into three tanksections is provided; that is, two tank sections 21 a and 21 b locatedon the downstream side of the air flow and one tank section 23 locatedon the upstream side of the air flow is provided.

Note that, at the lower end of the leftmost heat transfer plate 12adjacent to the left side end plate 24 having the coolant exit pipe 24 band the coolant inlet pipe 24 a, the tank section 21 a is solelyprovided and the tank section 21 b is not provided on the downstreamside of the air flow. A barrier wall for interrupting the coolantpassage (a blind lid structure having no communicating opening) isprovided at the position at which the tank section 21 b is not formed.

The coolant inlet pipe 24 a in the left end plate 24 communicates with aflow passage of the tank section 20 at the upper end of the heattransfer plate 12 on the downstream side of the air flow. In the flowpassage of this tank section 20, as no partition is arranged at anintermediate position in the stacking direction of the heat transferplates 12 (a boundary between the left side area X and the right sidearea Y), the flow passage of the tank section 20 passes throughout thelength thereof in the stacking direction of the heat transfer plates 12(in the leftward/rightward direction).

Accordingly, the coolant entering from the coolant inlet pipe 24 a flowsthrough the passage of the tank sections 20 along a total length in thestacking direction of the heat transfer plates 12. The coolant descendsthe air flow downstream side coolant passage 15 of the heat transferplate 12 as indicated by arrows n1 and n2. Here, the arrow n1 indicatesthe coolant descending the coolant passage 15 located in the left sidearea X, and the arrow n2 indicates the coolant descending the coolantpassage 15 located in the right side area Y.

The heat transfer plate 12 is constituted so that coolant passage 15 inthe left side area X communicates solely with the air flow downstreamside tank section 21 b at the lower end of the heat transfer plate 12,and the coolant passage 15 in the right side area Y communicates solelywith the air flow downstream side tank section 21 a.

The flow passage in the tank section 21 a communicates with the left endflow passage of the air flow upstream side lower tank section 23 via thecommunication passage 24 c formed in the vicinity of the lower end ofthe left side end plate 24.

In this flow passage of the lower tank section 23, a partition (notshown) is disposed at an intermediate position in the stacking directionof the heat transfer plates 12 (the boundary between the left side areaX and the right side area Y) to divide the flow passages in the leftside area X and the right side area Y. Accordingly, the communicationpassage 24 c communicates solely with the flow passage in the left sidearea X of the lower tank section 23.

On the other hand, the flow passage in the tank section 21 bcommunicates with the right end low passage of the air flow upstreamside lower tank section 23 via the communication passage 25 a formed inthe vicinity of the lower end of the right end plate 25. That is, thecommunication passage 25 a communicates solely with the flow passage inthe right side area Y among the flow passages in the lower tank section23. The coolant, descending as indicated by the arrow n1, flowsrightward through the lower tank section 21 b as indicated by an arrowp1, and then flows into the right side flow passage of the air flowupstream side lower tank section 23 via the communication passage 25 aof the right end plate 25 as indicated by an arrow q1.

The coolant in the right side flow passage of the lower tank section 23rises in the air flow upstream side coolant passage 16 in the right sidearea Y as indicated by an arrow r1, and flows into the right side flowpassage of the air flow upstream side upper tank 21.

On the other hand, the coolant descending the coolant passage 15 in theright side area Y located on the air flow downstream side, as indicatedby an arrow n2, flows leftward in the lower tank section 21 a asindicated by an arrow p2, and then flows into the left side flow passageof the air flow upstream side lower tank section 23 via thecommunication passage 24 c of the left end plate 24 as indicated by anarrow q2.

The coolant in the left side flow passage of the lower tank section 23rises the air flow upstream side coolant passage 16 in the left sidearea Y as indicated by an arrow r2, and flows into the left side flowpassage of the air flow upstream side upper tank 21.

The coolant coming from the coolant passage 16 in the right side area Yand the coolant coming from the coolant passage 16 in the left side areaX join together in the upper tank 21 and flow toward the coolant exitpipe 24 b as indicated by an arrow s.

Thereby, between the air flow downstream side upper tank section 20communicating with the coolant inlet pipe 24 a and the air flow upstreamside upper tank section 21 communicating with the coolant exit pipe 24b, a first coolant passage indicated by the arrows n1, p1, q1 and r1 anda second coolant passage indicated by the arrows n2, p2, q2 and r2 arearranged in parallel to each other.

In this regard, according to the inventive plate type evaporator 10, thefin sections 17 are arranged between the adjacent rib sections 14. Thus,when the space pitch Sp is enlarged for the purpose of restricting anincrease in ventilation resistance caused by the arrangement of the finsections 17, the number of heat transfer plates 12 to be stackedtogether decreases.

The reduction of the number of heat transfer plates 12 causes thereduction of the coolant passage area, which increases the pressure lossof the coolant passage in the evaporator 10. The increase in pressureloss of the coolant flow passage causes the rise of the coolantevaporation temperature, whereby the cooling performance of theevaporator 10 becomes worse.

In the first embodiment, as the coolant inlet pipe 24 a and the coolantexit pipe 24 b are coupled to each other by a single coolant passagearranged in series indicated by the arrows a to k, the above-mentionedincrease in pressure loss is liable to occur in the coolant passage.

Contrarily, in the coolant passage structure of the sixth embodiment, asthe first coolant passage and the second coolant passage are coupled inparallel to each other in the evaporator 10, it is possible toeffectively suppress the increase in the pressure loss in the evaporator10.

By coupling the first and second coolant passages in an X patternbetween the air flow downstream side upper tank section 20 and the airflow upstream side lower tank section 21, it is possible to make thedistribution of the air temperature blown out from the evaporatoruniform.

Seventh Embodiment

As shown in FIG. 4, according to the first embodiment, the fin section17 is a slit fin having a U-shape, but the slit fin is not limited tohave such a U-shape. The seventh embodiment relates to another shape ofthe slit fin constituting the fin section 17. As shown in FIG. 17, theslit fin constituting the fin section 17 is protruded to have a smoothlycurved surface (a dome-like contour).

According to the curved surface (the dome-like contour) of the slit finshown in FIG. 17, the offset wall surface 17 a and the left and rightside walls 17 b and 17 c are continuously coupled by a smooth curve.

Eighth Embodiment

The width dimension Fw of the fin section 17 constituted by the slit finis sufficiently smaller than the rib pitch Rp; in other words, a widthdimension of a flat surface of the base plate section 13; in the firstand second embodiments as shown in FIGS. 3 and 9. Contrarily, in theeighth embodiment, as shown in FIG. 18, a width dimension Fw of the finsection 17 constituted by the slit fin is sufficiently larger than inthe first embodiment.

In the eighth embodiment, similarly to the second embodiment, the ribsections 14 in the two heat transfer plates 12 are arranged at the sameposition in the air flowing direction A. In this structure, the finsection (slit fin) 17 is formed to have a fin width dimension Fw nearlyequal to a width dimension (a dimension of the flat surface in the airflowing direction) of the flat surface of the base plate section 13located between the rib sections 14.

Concrete dimensions in the eighth embodiment are as follows; the spacepitch Sp (=Tp−2t): 3.0 mm; the thickness t of the heat transfer plate12: 0.15 mm; the rib section pitch Rp: 7.1 mm; the height Rh of the ribsection: 1.45 mm; the fin pitch Fp=the rib section pitch Rp; the finwidth Fw: 4.0 mm; and the fin height Fh: 1.0 mm.

According to the eighth embodiment, it is possible to increase the heattransfer area because the fin width Fw is enlarged, from (0.8 mm×2) inthe first embodiment, to 4.0 mm.

Ninth Embodiment

In a ninth embodiment, the space pitch Sp which is a mutual distancebetween the base plate sections 13 (that is, flat surface portions) ofthe heat transfer plates 12 adjacent to each other while interposing theair passage 18, is studied.

As shown in FIG. 19, in the ninth embodiment, the rib height Rha of therib section 14 is equal to a height from a surface of the base platesection 13 of the heat transfer plate 12 (that is, a height projectedinto the air passage 18). Accordingly, the projected height Rha is avalue obtained by subtracting the thickness t of the heat transfer plate12 from the rib height Rh in FIGS. 6 and 18 (Rha=Rh−t).

If the space pitch Sp becomes larger, a gap G between the rib sections14 projected into the air passage 18 increases, whereby the action ofthe rib section 14 for guiding the air stream becomes insufficient and,finally, the air stream linearly flows through the air passage 18.

The present inventors have specifically studied the relationship betweenthe space pitch Sp and the projected height Rha, and found that themeandering stream A1 is certainly formed by determining the space pitchSp to be three times the projected height Rha or less, that is, thespace pitch Sp≦3×Rha. Thereby, it has been confirmed that the dwellingregion F of the air stream along the surface of the base plate section13 (see FIG. 27) can be eliminated.

In this regard, as the pressure loss in the air stream increases if thespace pitch Sp is extremely small, the space pitch Sp must be larger bya predetermined amount than the rib height Rha of the rib section 14.Preferably, the space pitch Sp is selected within a range from Sp=(2.0to 2.3)×Rha for the purpose of forming the meandering air stream as wellas reducing the pressure loss of the air stream.

Tenth Embodiment

A tenth embodiment relates to the projected height Fha of the finsection 17 when the fin section 17 is constituted by the slit fin.

As shown in FIG. 20, according to the tenth embodiment, a projectedheight Fha of the fin section 17 is equal to a height from the baseplate section 13 of the heat transfer plate 12 (that is, a projectedheight into the air passage 18). More concretely, the projected heightFha is equal to a distance between the surface of the base plate section13 of the heat transfer plate 12 and a center of a thickness of theoffset wall surface 17 a. Therefore, the projected height Fha of the finsection 17 is a value obtained by subtracting a thickness t of the heattransfer plate 12 and half a thickness t′ of the offset wall surface 17a from the fin height Fh, that is Fha=Fh−t−0.5t′.

On the other hand, an axis H of a plate is parallel to the base platesection 13 of the heat transfer plate 12 (see FIG. 20). A perpendicularline I orthogonal to the axis H of the plate is drawn. A length of aline is defined as L, which connects intersecting points J1 and J2 toeach other, of the perpendicular line I on the surfaces of the two heattransfer plates 12 opposed to each other while interposing the airpassage 18. The projected height Fha of the fin section 17 is determinedto be half the length L or less at a position at which the fin section17 is formed. That is, Fha≦0.5×L.

In FIG. 20, the fin section 17 at a position a is disposed to oppose thebase plate section 13 in the adjacent heat transfer plate 12, the finsection 17 at a position b is disposed to oppose the front (top) of therib section 14 in the adjacent heat transfer plate 12, and the finsection 17 at a position c is disposed to oppose an intermediate heightportion between a top and a root of the curved surface in the adjacentheat transfer plate 12.

Accordingly, the lengths defined as described above have therelationship of L1>L3>L2. In either of the fin section 17 in a, b or c,the projected height Fha1, Fha2 or Fha3 is half a line length L1, L2 orL3 or less.

That is, the following relationship is established; Fha1≦0.5×L1,Fha2≦0.5×L2, and Fha3≦0.5×L3.

The line lengths L1, L2 and L3 are plate gaps determining thecross-sectional area of the passage variously changing in accordancewith the directions A of air stream in the air passage 18 formed betweenthe adjacent two heat transfer plates 12.

Thus, by setting the projected heights Fha1, Fha2 and Fha3 of the finsection 17 as described above, even if the position of the fin section17 varies to the position a, b or c, a position of a center of a platethickness of the offset wall surface 17 a in the fin section 17 isalways located closer to the base plate section 13 (the base platesection 13 n which the fin section 17 is formed) than to the center ofthe above-mentioned “plate gap determining the cross-sectional area ofthe air passage”.

As the offset wall surface 17 a in the fin section 17 is located in theair stream in the air passage 18 and extends parallel to the flatsurface of the base plate section 13 (parallel to the air flowingdirection A), the air is liable to flow along the offset wall surface 17a.

Therefore, as the offset wall surface 17 a is located closer to the baseplate section 13 than to the center of the “plate gap determining thecross-sectional area of the air passage”, it is possible to cause theair stream flowing along the offset wall surface 17 a to approach thebase plate section 13. As a result, as shown in FIG. 21, it is possibleto assuredly form the air stream A1 largely meandering closer to thebase plate section 17 than to the top of the curved surface of the ribsection 14. Thus, the dwelling region F (see FIGS. 22 and 27) of the airstream flowing along the surface of the base plate section 13 iseliminated.

Contrarily, when the projected height Fha of the fin section 17 is toohigh, that is, when the projected height Fha of the fin section 17 islarger than the above-mentioned line length L, the offset wall surface17 a of the fin section 17 is closer to the top of the opposed ribsection 14 as shown in FIG. 22, whereby the air stream flowing along theoffset wall surface 17 a is away from the base plate section 13 and,instead, approaches the top of the rib section 14.

In other words, according to the comparative example shown in FIG. 22,the offset wall surface 17 a disturbs the formation of the meanderingstream to be established by the rib section 14. As a result, the airstream becomes almost linear as indicated by an arrow A2, causing thedwelling region F of the air stream along the surface of the base platesection 13, and the extreme reduction of the heat transfer rate on thesurface of the base plate section 13.

In this regard, if the projected height Fha of the fin section 17 isextremely small, it is difficult to pass air through the interior of theoffset wall surface 17 a, whereby it is necessary that the projectedheight Fha of the fin section 17 is a predetermined height or morecapable of ensuring the air stream within the interior of the offsetwall surface 17 a.

According to the tenth embodiment, while the projected height Fha of thefin section 17 is set to be half a line length L or less at a positionat which the fin section 17 is formed; that is, Fha≦0.5×L, it isinevitable that the projected height Fha of the fin section 17 has theproduction variance (machining tolerance) when the heat exchanger ismanufactured. Concretely, the machining tolerance is usuallyapproximately ±17%, and if the projected height Fha is, for example, 3mm or less, the projected height Fha of the fin section 17 hasapproximately ±0.5 mm.

Accordingly, “to suppress the projected height Fha of the fin section 17to a value half a line length L or less at the position at which the finsection 17 is formed” does not strictly mean that the height Fha must behalf a line length L or less, but means that it is generally half a linelength or less including the excess amount due to the above-mentionedmachining tolerance.

Eleventh Embodiment

In the above-mentioned embodiments, the offset wall surface 17 a of thefin section 17 is formed to be parallel to the flat surface of the baseplate section 13. Contrarily, in an eleventh embodiment, the offset wallsurface 17 a of the fin section 17 is inclined to the flat surface ofthe base plate section 13.

As shown in FIG. 23, according to the eleventh embodiment, when the finsections 17 are located adjacent to each other both on windward andleeward sides of the rib section 14, the offset wall surface 17 a of thefin section 17 is inclined in the same direction as the curved surfaceof the closest rib section 14 on the same heat transfer plate 12.

That is, the offset wall surface 17 a of the fin section 17 located onthe windward side of the rib section 14 is inclined to be away from theflat surface of the base plate section 13 as going from the upstream tothe downstream. Contrarily, the offset wall surface 17 a of the finsection 17 located on the leeward side of the rib section 14 is inclinedto be closer to the flat surface of the base plate section 13 as goingfrom the upstream to the downstream.

Thereby, the offset wall surface 17 a of the fin section 17 performs theoperation for facilitating the guiding of the air stream due to thecurved surface of the rib section 14 (the guiding operation for themeandering stream). As a result, as shown in FIG. 23, the meanderingstream A3 is assuredly formed, whereby the region F in which the airstream dwells along the surface of the base plate section 13 (see FIGS.22 and 27 is eliminated.

Twelfth Embodiment

The fin sections 17 are arranged adjacent to the rib section 14 both onthe windward and leeward sides in the eleventh embodiment. Contrarily,as shown in FIG. 24, according to a twelfth embodiment, the fin section17 is disposed adjacent to the rib section 14 solely on the leewardside, so that the offset wall surface 17 a of the fin section 17 isinclined in the same direction as the leeward side curved surface of therib section 14. That is, the offset wall surface 17 a is inclined closerto the flat surface of the base plate section 13 from upstream todownstream.

FIG. 25A illustrates a comparative example wherein the fin sections 17are not disposed both on the windward and leeward sides of the ribsection 14. As the air stream indicated by an arrow K impinges to thewindward curved surface of the rib section 14, the heat transfer ratebecomes high. However, vortices are generated on the leeward side of therib section 14 as indicated by an arrow M due to the impingement on thewindward side indicated by the arrow K to result in the dwelling of theair stream.

As a result, the heat transfer rate on the leeward side curved surfaceof the rib section 14, disposed in the region in which the vortices Mare generated, becomes extremely low. Similarly, also in the base platesection 13, the heat transfer rate extremely deteriorates in the regionin which the vortices M are generated. In this regard, O in FIG. 25Adenotes a position at which the air stream impinges again to the baseplate section 13. In a portion of the base plate section 13 upstreamfrom the position O, the heat transfer rate is low.

Contrarily, in the twelfth embodiment, as shown in FIG. 25B, it ispossible to cause the air stream P passing through the inside of theoffset wall surface 17 a of the fin section 17 to flow closer to theleeward side curved surface of the rib section 4.

Thus, as the region in which the vortices M′ generate (the air streamdwelling region) can be minimized to a great extent in comparison withthe region in which the vortices M are generated, it is possible tolargely improve the heat transfer rate of the leeward side curvedsurface of the rib section 14 and the base plate section 13.

In this regard, in the above-mentioned eleventh and twelfth embodiments,when the offset wall surface 17 a of the fin section 17 is disposeddirectly adjacent to the rib section 14, the offset wall surface 17 a ofthe fin section 17 is inclined in the same direction as in the curvedsurface of the closest rib section 14. However, as shown in FIGS. 19 and21, when the fin section 17 is disposed at a center of the base platesection 17 as seen in the air flowing direction and the fin section 17is disposed opposite to a front of the rib section 14 in the oppositeside heat transfer plate 12, it is better to form the offset wallsurface 17 a of the fin section 17 parallel to the base plate section13, instead of inclining the same.

That is, if the fin section 17 is disposed opposite to a front of therib section 14 of the opposite side heat transfer plate 12, the offsetwall surface 17 a is just located at a position at which the air streamis reversed. Accordingly, if the offset wall surface 17 a is inclined,air flows along the inclination thereof to disturb the reverse of theair stream. Contrarily, if the offset wall surface 17 a is formedparallel to the base plate section 13, the offset wall surface isneutral against the air stream, whereby the offset wall surface 17 adoes not guide the air stream to a particularly inclined direction.Thus, as shown in FIGS. 19 and 21, the air stream is smoothly reversedand the meandering stream A1 is suitably formed.

Thirteenth Embodiment

A thirteenth embodiment relates to the determination of dimensionsbetween the offset wall surface 17 a of the fin section 17 and thesurface of the heat transfer plate 12.

When the heat exchanger is a cooler type generating condensation wateras air is being cool, such as the air conditioner evaporator 10 shown inFIG. 1, the drainage of condensation water is an important problem inthe design of the heat exchanger.

According to the thirteen embodiment, gaps Q1, Q2 are determined to be0.3 mm or more between the offset wall surface 17 a of the fin section17 and the surface of the heat transfer plate 12; more concretely, a gapQ1 between the inside surface of the offset wall surface 17 a and thesurface of the base plate section 13, and a gap Q2 between the outsidesurface of the offset wall surface 17 a and the surface of the ribsection on the heat transfer plate 12 disposed on the opposite sideshown in FIG. 26.

According to a study made by the present inventors, it has beenconfirmed that by setting the above-mentioned gaps Q1 and Q2 to 0.3 mmor more (Q1, Q2≧0.3 mm), the condensation water does not block thesegaps Q1 and Q2 but is smoothly drained.

In the cooler type heat exchanger generating condensation water, theinstallation posture of the heat exchanger during the use is determinedso that the longitudinal direction of the rib section 14 (vertical to apaper surface in FIG. 26) coincides with the direction of gravity(upward/downward direction). Thereby, the condensation water generatedon the surface of the heat transfer plate 12 smoothly flows down in thelongitudinal direction of the rib section 12.

Other Embodiments

In the above-mentioned embodiments, a description has been made for thecases wherein the coolant passages (internal passages) 15 and 16 areformed inside the rib section 14 by laying and fixing two heat transferplates 12 completely separated from each other. As disclosed in FIG. 36of Japanese Unexamined Patent Publication No. 2001-41678, however, thetwo heat transfer plates 12 and 12 constituting the coolant passages(internal passages) 15, 16 may be formed of a press-formed single platemember which is bent at a widthwise center to be two sections 12, 12,and thereafter base plate sections 13, 13 thereof are fixed together toform the coolant passages 15 and 16.

Further, lateral surfaces of the respective plate members constitutingthe above-mentioned two heat transfer plates 12, 12 may be coupledtogether with a cleat-like coupler. This coupler is designed to have thesame length as the space pitch Sp. Such a coupling structure is alsodisclosed in FIG. 36 of Japanese Unexamined Patent Publication No.2001-41678.

As understood from such modifications, “two heat transfer plates 12 areused as one pair” in the present invention includes both of a casewherein the completely separated two heat transfer plates 12 are stackedtogether and another case wherein a single plate member 120 is bent at acenter 121 and two portions of half a size are laid together.

In the above embodiments, while the description has been made on a casewherein the present invention is applied to an evaporator 10 which is aheat-suction side heat exchanger for a refrigeration cycle, the presentinvention may be applicable to heat exchangers for various uses.

For example, the present invention may be applicable to a condenserwhich is a heat-radiation side heat exchanger for a refrigeration cycle.Also, the present invention may be applicable to a heat exchangerwherein hot water flows through the internal passage of the heattransfer plate 12 (the coolant passages 15 and 16 in the above-mentionedembodiments) such as a hot water type radiator for a heater or aradiator for cooling an engine.

Similarly, the present invention may be applicable to a heat exchangersuch as an engine oil cooler wherein oil flows through internal passagesor an heat exchanger wherein cold water flows through internal passages.

While the invention has been described by reference to specificembodiments chosen for purposes of illustration, it should be apparentthat numerous modifications could be made thereto by those skilled inthe art without departing from the basic concept and scope of theinvention.

1. A heat exchanger wherein a plurality of heat transfer plates formingplate surfaces extending in the flowing direction (A) of external fluidare stacked together, a gap is provided between said plate surfaces ofsaid adjacent heat transfer plates to form an external passage throughwhich said external fluid flows, a plurality of rib sections extendingorthogonal to the flowing direction A of said external fluid areprojected from said plate surfaces into said external passage to beintegral with said heat transfer plates, by shifting positions of theplurality of rib sections in one of said adjacent heat transfer platesrelative to positions of the plurality of rib sections in the other ofsaid adjacent heat transfer plates as seen in the flowing direction A ofsaid external fluid, said external passage is formed in a meanderingmanner, the plurality of rib sections form an internal passage insidethereof, through which flows internal fluid, fin sections are projectedfrom said plate surfaces at positions between the adjacent rib sectionsto be integral with the heat transfer plate, and said fin section ispress-formed so as to protrude a cut portion partially cut a platethickness of said heat transfer plate.
 2. A heat exchanger as defined byclaim 1, wherein said heat transfer plates are combined to form pairs,and said rib sections and said fin sections are formed integral withsaid pair of heat transfer plates, and the pair of heat transfer platesare fixed together to form said internal passage inside the plurality ofrib sections.
 3. A heat exchanger as defined by claim 2, whereinpositions in the pair of heat transfer plates at which said rib sectionsare formed are shifted in the flow direction A of the external fluid,and said internal passage is formed by said rib sections formed in oneof the pair of heat transfer plates and a plate surface of the other. 4.A heat exchanger as defined by claim 2, wherein said rib sections areformed in said pair of heat transfer plates at the same positions asseen in the flowing direction A of said external fluid, and saidinternal passages are formed by the combination of said rib sectionsformed in said pair of heat transfer plates, respectively.
 5. A heatexchanger as defined by claim 1, wherein said heat transfer plate isconstituted by a single extrusion-formed plate material, said ribsections are formed by extrusion-forming a tubular-shaped portion onsaid single extrusion-formed plate material, and said fin sections areformed integral with said single extrusion-formed plate material to beprojected from a plate surface of said single extrusion-formed platematerial.
 6. A heat exchanger as defined by claim 1, wherein said heattransfer plate has a base plate section having a flat surface betweenthe adjacent rib sections, and said fin section is formed in said baseplate section.
 7. A heat exchanger as defined by claim 1, wherein awidth (Fw) in the flowing direction (A) of said external fluid of saidfin section is 5 mm or less.
 8. A heat exchanger as defined by claim 1,wherein said fin section is a slit fin having an offset wall surfaceapart from a plate surface of said heat transfer plate at apredetermined gap, wherein said offset wall surface are coupled to aplate surface of said heat transfer plate at two positions.
 9. A heatexchanger as defined by claim 8 wherein, when a gap between positions onthe pair of heat transfer plates opposed to each other to define saidexternal passage, at which positions are formed said slit fins, isdefined as L, and a projected height of said offset wall surface from aplate surface of said heat transfer plate is defined as Fha, thefollowing relationship is satisfied:Fha≦½L.
 10. A heat exchanger as defined by claim 8, wherein across-sectional shape of said rib section has a curved surface projectedfrom the surface of said heat transfer plate, which is generallysemicircular, said slit fin is located at a position directly ondownstream side from said external fluid relative to said rib section,and said offset wall surface is inclined in the same direction as theinclination of the downstream side curved surface in the generallysemicircular curved surface of said rib section.
 11. A heat exchanger asdefined by claim 8, wherein the cross-sectional shape of said ribsection is such that it has a curved surface protruded semi-circularlyfrom a surface of said heat transfer plate, said slit fin is disposedadjacent to said rib section at a position directly on the upstream sideof said external fluid, and said offset wall surface is inclined in thesame direction as the inclination of the upstream side curved surface ina generally semicircular curved surface of said rib section.
 12. A heatexchanger as defined by claim 8, wherein said slit fin is disposedopposite to a front of said rib section while interposing said externalpassage, and said offset wall surface is formed to be parallel to a flatsurface of said heat transfer plate.
 13. A heat exchanger as defined byclaim 8, wherein said external fluid is air and said internal fluid is acoolant for cooling said air, wherein said heat exchanger is constitutedas a cooling heat exchanger generating condensation water on the surfaceof said heat transfer plate, and a gap (Q1, Q2) between said offset wallsurface and the surface of said heat transfer plate is 0.3 mm or more.14. A heat exchanger as defined by claim 1, wherein said fin section isa protruded fin having a predetermined angle relative to the surface ofsaid heat transfer plate.
 15. A heat exchanger as defined by claim 14,wherein said protruded fin is triangular.
 16. A heat exchanger asdefined by claim 15, wherein said triangular protruded fin is inclinedto the flowing direction (A) of said external fluid at an angle from 15°to 45°.
 17. A heat exchanger as defined by claim 14, wherein saidprotruded fin is rectangular.
 18. A heat exchanger as defined by claim14, wherein the inclination angle of said protruded fin relative to theflowing direction (A) of said external fluid is determined in a smallangle range from −30° to +30° so that a surface of the protruded finfollows the flow direction (A) of said external fluid.
 19. A heatexchanger as defined by claim 14, wherein said external fluid is air andinternal fluid for cooling said air flows through said internal passage,said heat transfer plate is disposed so that the longitudinal directionof said rib section coincides with the upward/downward direction, and aninclination angle of said protruded fin is in a range from 60° to 120°relative to the flowing direction (A) of said external fluid so that asurface of said protruded fin follows the longitudinal direction of saidrib section.
 20. A heat exchanger as defined by claim 1, wherein saidinternal passage has an upstream side internal passage disposed on theupstream side in the flow direction (A) of said external fluid and adownstream side internal passage disposed on the downstream side in theflowing direction (A) of said external fluid, said upstream sideinternal passage and said downstream side internal passage arerespectively sectioned vertically to the flowing direction (A) of saidexternal fluid into a plurality of areas (X, Y), and passages connectedin parallel to each other are constituted between the plurality of areas(X, Y) of said upstream side internal passages and the plurality ofareas (X, Y) of said downstream side internal passages.
 21. A heatexchanger as defined by claim 20, wherein said downstream side internalpassage is an inlet side passage for said internal fluid, and saidupstream side internal passage is an exit side passage for said internalfluid.
 22. A heat exchanger as defined by claim 20, wherein saidparallel passages couple the plurality of areas (X, Y) in said upstreamside internal passage and the plurality of areas (X, Y) in saiddownstream side internal passage to each other in an X pattern.